This invention concerns an improved apparatus and methods for cooling and sealing the compressed gas in a rotary helical screw compressor using any type of gas, whether or not the gas is highly superheated at suction pressure conditions, and whether or not the gas is highly soluble in the compressor oil, to optimize the effectiveness of the compressor oil in both cooling the gas and sealing the rotor edges and to maximize both the isothermal and volumetric efficiencies of the gas compression process. As noted in the prior art, a lubricating fluid such as a hydrocarbon oil is incorporated within and circulated through a refrigeration or gas compression circuit utilizing a helical screw rotary compressor to compress the working fluid. The lubricating oil performs multiple functions, one of which is to lubricate the moving parts of the compressor, such as the bearings and seals. The same oil is also used to seal the compression chamber defined by the moving parts, i.e., the intermeshed helical screw rotors within the casing bores during their rotation, and at the same time it is used to cool the working fluid. The compression raises the temperature of the working fluid, so that both the working fluid itself and the lubricating oil must be cooled upon discharge from the compression chamber. Conventionally, oil that is miscible with the refrigerant or mixed with the gas is discharged with the working fluid at a high pressure from the compressor, is separated from the working fluid in an oil separator, and returned to the compressor. Typically, the oil is cooled within an oil cooler and is pressurized by an oil pump prior to injection into the compressor via one or more injection ports opening to the compression process itself. The injection port for the oil intended for sealing is typically the very same one used to inject the oil intended for cooling so that there is no distinction between the location of the injection port or ports for the oil used for cooling the gas or sealing the clearance spaces or lubricating the rotors. In the case of refrigerant gases, oftentimes, to eliminate the oil cooler, refrigerant in liquid form is diverted from the refrigeration cycle and injected via one or more ports either opening to the compression process itself near the discharge end of the rotors or, following the compression process, opening to the discharge port of the compressor. In either case, the temperature of the gas and oil mixture at the discharge of the compressor is lowered to the level equivalent to that obtained by the separate oil cooler, the oil cooler being cooled typically either by liquid refrigerant diverted from the refrigeration cycle or by water. The injection of liquid refrigerant to the compression process itself is referred to in the industry as Liquid Injection.
As far back as 1962, Nilsson and Wahlsten proposed, in Canadian patent 643,525, to improve the cooling of the working fluid by providing the liquid, typically a lubricating oil but possibly other liquids such as water, in very finely divided form through a series of holes at various locations in the compressor casing. Such holes were shown distributed along the upper cusp of the compressor casing and also in the suction port area in close proximity to the suction side ends of the rotors. The holes in the suction port area direct the liquid along the axis of rotation of the rotors and face the suction side ends of the rotors. They also proposed that the rotors themselves be made hollow and therefore capable of conducting the liquid out through atomizing holes that lead directly into the gas compression pockets formed by the intermeshing of the male and female rotors.
In 1966, in U.S. Pat. No. 3,265,293, Schibbye disclosed a rotary screw compressor acting as a vacuum pump in which, as he noted is old in the art, liquid is introduced into the working space of the compressor to aid in sealing the running clearance spaces and for directly cooling the contents of the compression chambers to reduce the temperature rise thereof as the work of compression is done thereon. Schibbye illustrates the introduction of such liquid by a supply pipe delivering a spray of liquid into the compressor intake. The end of the supply pipe is suspended within the suction intake. The liquid is introduced solely through the supply pipe and for the dual purpose of sealing the running clearance spaces and directly cooling the contents of the compression chambers. Schibbye noted also that it will be understood that other and equivalent means for introducing liquid into the compressor, such as that disclosed by Nilsson and Wahlsten in U.S. Pat. No. 3,129,877, may be employed.
A design similar to that of Nilsson and Wahlsten in Canadian Patent 643,525, showing nozzles in the suction port area in close proximity to the suction side ends of the rotors, the nozzles mounted in the compressor casing, was presented by Shaw in 1985 in U.S. Pat. No. 4,497,185. In this design, all of the oil intended for cooling and sealing the working fluid is atomized at the end plates of the compressor on the suction side. The nozzles themselves are mounted in the compressor casing facing the inlet end of the intermeshed helical screw rotors. An alternative location is presented wherein the nozzles are mounted on the compressor casing perpendicular to the rotor axes at a point just after the gas or refrigerant suction charge is locked in the rotors at a closed thread. This alternative is proposed when the gas or refrigerant is highly soluble in the oil.
In 1974, Zweifel, in U.S. Pat. No. 3,820,923, disclosed an apparatus whereby oil is atomized and injected through approximately 100 very small holes drilled in the compressor casing circumferentially around near the discharge end of the rotors.
It is of interest to note that Nilsson and Wahlsten, in U.S. Pat. No. 3,129,877, which was issued in 1964, state that it is highly desirable that compression be commenced without preheating of the inlet air and that by confining the introduction of liquid to or approximately to the compression phase of the cycle, undesirable preheating of the inlet air by recirculated liquid at higher than inlet temperature is with certainty avoided.
For simplicity in disclosing the present invention, the lubricating oil or other liquid such as water or refrigerant in liquid form which is used for lubrication or sealing or cooling will be referred to as the nonworking liquid. The compressed gas, vapor or refrigerant will be referred to as the working fluid.
There are two disadvantages to the atomization process when the working fluid is a refrigerant such as R-12 or R-22 that is highly soluble in the nonworking liquid, i.e., the injection of atomized oil at the suction port at a temperature in the range of 50.degree. C. into the working fluid that may be as cold as -35.degree. C. could cause heating and expansion of the working fluid prior to entering the compression chamber. Furthermore, the injection into the working fluid at the suction port of atomized oil from the discharge side of the oil separator sump could liberate significant quantities of dissolved working fluid into the suction side prior to entering the compression chamber defined by the rotors and casing of the compressor. In both cases, the volumetric efficiency of the compression would decrease.
In addition, depending upon the geometrical relationship of the suction port to the rotors, mounting the nozzles within the compressor casing, as specified in the prior art, can cause the nonworking liquid oil flow to be transverse to the working fluid gas flow, thereby diminishing the probability of a homogeneous mixture entering the compression chamber and increasing the tendency for the oil droplets to accumulate on the inner surfaces of the suction intake port of the compressor.
Most attempts to improve the efficiency of the rotary screw compressor have been oriented towards improving the effectiveness of the oil injection system. However, it is also possible to improve compressor efficiency by providing more than two rotors within the same casing, therein reducing the volume of the clearance space between the tips of the rotors and the compressor casing with respect to the volumetric flow rate capacity of the compressor. However, in the prior art, disclosures of screw compressors in which the casing houses more than two rotors do not indicate any attempt at reducing the volume of the clearance space between the tips of the rotors and the compressor casing with respect to the volumetric flow rate capacity of the compressor.
For example, in 1963, Bailey, in U.S. Pat. No. 3,073,513, indicates as an objective to provide a rotary compressor of the positive displacement type including two or more rotors disposed within a housing and formed with intermeshing helical lobes and grooves, which, however, are not in physical contact with one another, but engage with small clearances, in which a liquid is introduced into the compressor in sufficient amounts to seal the clearances and also to enable one rotor to drive the other or others without the necessity for the usual intermeshing timing gears hitherto employed. However, no further spatial relationship between the rotors is described other than to show the conventional single male and single female intermeshing rotors.
In 1964, in U.S. Pat. No. 3,133,695, Zimmern introduced what is known in the industry as the "Monoscrew" compressor, but which actually consists of three rotors within the same housing. In the center is an hourglass-shaped screw rotor which is flanked by two intersecting "gate" or worm gear rotors whose axes of rotation are perpendicular to the central hourglass rotor. This type of compressor is considered in the art to be a totally separate category of rotary screw compressor, and therefore is not germane to the objective of reducing the volume of the rotor to casing clearance space with respect to the volumetric flow rate capacity of the dual screw compressor.
In 1976, in Federal Republic of Germany Patent P26 21 303.6-15, Maekawa disclosed a screw compressor unit in which two axially adjacent sets of rotatable screws are mounted within the same housing, the first rotors and the second rotors being coaxially interconnectable via first and second shafts. In effect, this compressor consists of two sets of male and female intermeshing screw rotors within a single housing, the sets of rotors being longitudinally separated by the first and second shafts. Again, there is no attempt at reducing the volume of the clearance space between the tips of the rotors and the compressor casing with respect to the volumetric flow rate capacity of the compressor.